Fluid dampened support having variable stiffness and damping

ABSTRACT

A fluid dampened support for a bearing such as a tilt pad bearing or a rolling element bearing such as a ball bearing, needle bearing, roller bearing and the like. The fluid dampened support includes a network of closely spaced beams which act as structural springs which support the outer race of the roller bearing for movement in any direction. The spring rate or constant of these structural springs can be caused to change after a selected amount of deflection. The amount of deflection needed to cause the change in spring rate can be adjusted to suit the particular application. A liquid is provided in the spaces between the beams to dampen movement of the pads. The damping rate can be made to change with movement of the pads. The structure and space between the beams is designed so as to provide virtually any reasonable damping characteristics. By providing such damping characteristics, the bearings are able to limit or damp out the vibrations occurring when the supported member passes through natural frequencies before reaching operating speed. This allows use of the rolling element bearing in high speed equipment such as compressors and turbines.

This application is a continuation, of application Ser. No. 07/945,694filed Sep. 16, 1992, abandoned, a CIP of Ser. No. 07/876,718, Apr. 24,1992, abandoned, a CIP of Ser. No. 07/309,081, Feb. 8, 1989, U.S. Pat.No. 5,137,373, a CIP of Ser. No. 07/283,529, Oct. 25, 1988, U.S. Pat.No. 5,112,143, a CIP of Ser. No. 07/055,340, May 29, 1987, abandoned,and a CIP of Ser. No. 07/878,601, May 5, 1992, U.S. Pat. No. 5,489,155,a CIP of Ser. No. 07/685,148, Apr. 15, 1991, abandoned, a CIP of Ser.No. 07/309,081, Feb. 8, 1989, U.S. Pat. No. 5,137,373.

FIELD OF THE INVENTION

The present invention relates to bearings and, more particularly, fluiddampened rolling element bearings and tilt pad bearings.

BACKGROUND OF THE INVENTION

Rolling element bearings such as ball, roller and needle bearings areused in almost every kind of machine and device with rotating parts.They are currently the most widely used bearing.

Rolling element bearings typically include four parts: an inner ring, anouter ring, the balls or rollers and a cage or separator for separatingthe balls from one another. The balls in ball bearings are normally madeof high carbon chromium steel. The balls are heat treated to highstrength and hardness and the surfaces are ground and polished.Cylindrical roller bearings are usually made of case hardened steel.

Rolling element bearings are made in a wide variety of types and sizes.Regardless of their size or shape, rolling element bearings operate onthe same basic principle of allowing low friction rotation of one memberrelative to the other while maintaining solid metal-to-metal contactbetween the two elements.

For a rotating shaft, relative rotation between shaft and bearing isusually prevented by mounting the inner ring with a press fit andsecuring it with a nut threaded on the shaft. Excessive interference ofmetal must be avoided in press-fits, or the stretching of the inner ringmay decrease the small but necessary internal looseness of the bearing.

Conventionally, rolling element bearings are mounted to a fixed housingso that because there is little radial play. Although the outer ring,when the shaft rotates, is mounted more loosely than the inner ring,rotational creep between the ring and the housing should be prevented.

Compared to other bearings such as conventional journal bearings,rolling element bearings offer a number of advantages. These include lowstarting friction; the ability to support loads inclined at any angle inthe transverse plane; the ability to support thrust components of loads;and low maintenance cost. In addition the bearings are easily replacedwhen worn out and require less axial space than for journal bearings.

There are, however, certain disadvantages associated with conventionalrolling element bearing assemblies. The cost is typically higher, moreradial space is generally required than with journal bearings and morenoise is generated by ball bearings, especially after wear. In addition,rolling element bearings are more subject to fatigue failure and aremore easily damaged by foreign matter. All rolling element bearings havea limited life, typically less than 20,000 hours depending on theapplication. Another disadvantage associated with rolling elementbearings is that they have very little damping capability because of themetal-to-metal contact between elements. Thus, rolling element bearingsare typically less well suited to overload and shock conditions. This isa significant drawback in high speed turbo machinery.

High speed equipment such as the compressor turbine in a jet engine andaeroderivative applications such as steam turbines, gas turbines andcompressors must pass through several natural frequencies beforereaching operating speed. When a system operates at its naturalfrequency or resonance, the system/rotor vibration amplitudes becomelarge. These vibrations can be destructive or even catastrophic if notadequately dampened. Bearings with adequate damping characteristicslimit or damp out the vibrations to allow the equipment to safely passthrough the critical speeds. Likewise, smaller vibrations due tounbalance can be dampened by the bearings damping characteristics.

As previously noted, rolling element bearings where metal-to-metalcontact exists have very little damping capability. Accordingly, whenrolling element bearings are used in jet engines or other high speedturbo machinery they must be supported in complex expensive multipartassemblies which use a squirrel cage centering spring. Examples of suchconstructions are shown in the following U.S. patents: U.S. Pat. No.3,456,992 to Kulina disclosing fluid retained between sealing rings;U.S. Pat. No. 3,863,996 to Raimondi disclosing a fluid dampened journalbearing; U.S. Pat. No. 3,994,541 to Geary et al. disclosing a fluiddampened tilt pad bearing; U.S. Pat. No. 4,097,094 to Gardner disclosinga fluid dampened pad-type bearing and U.S. Pat. No. 4,213,661 to Marmoldisclosing an O-ring type damper. Another form of damper was recentlyproposed by Messrs. Heshmat and Walton of Mechanical Technology Inc.These so-called multi-squeeze film dampers use a spiral foil to providea spiral multi-film damper.

There are a number of disadvantages associated with known squeeze filmdamper bearings. Squeeze film dampers which use a squirrel cagecentering spring typically occupy an axial space 2 to 3 times largerthan the axial space available for the squeeze film land. Moreover, itis very difficult to install the centering spring and center the rotorwithin the squeeze film clearance. For this reason, the performance ofthe damper is often not consistent from one engine to another. Themulti-piece design and precision required to assemble such an element isalso very undesirable.

In process type compressors, elastomer O-rings are used as a centeringspring element in addition to providing sealing at the damper ends. Theelastomer rings are not reliable as spring elements and have a verynarrow range of stiffness. They degrade with time and temperature.Centering the damper with the O-rings is also difficult because theytend to creep due to the static loading. O-rings are also not capable oftaking any thrust load which is required in certain applications.

Another problem experienced in most conventional damper bearings iscavitation and air ingestion caused by negative pressure in the squeezefilm cavity. Such cavitation is a primary cause of poor performance ofconventional damper bearings.

Fluid film bearings, on the other hand, have significant dampingcapability from the fluid film. Of the available fluid film bearings,the so-called tilt-pad radial bearing is by far the mostuniversally-prescribed design for machines requiring maximumrotordynamic stability because of its exceptional stabilitycharacteristics. Consequently, it has become the standard by which manyother radial bearings are measured when seeking a highly stable bearingdesign. The tilt-pad bearing's popularity is evidenced by the largenumber of applications found in industry, both as original equipment,and as aftermarket replacements. Applications range from smallhigh-speed machines such as turbochargers and compressors, to very largeequipment such as steam turbines and generators.

The high rotordynamic stability comes from the reduction ofcross-coupled stiffness that occurs when pads are free to tilt abouttheir individual pivot points. This attenuates the destabilizingtangential oil film forces that can induce catastrophic subsynchronousvibration in machines equipped with conventional fixed-geometrybearings. Since so many machines are susceptible to this type ofbearing-induced instability, there is a large demand for qualitytilt-pad bearings.

Because of its many moving parts and manufacturing tolerances, thetilt-pad design is also the most complex and difficult to manufacture ofall journal bearing designs. The design complexity is evident in thenumber of highly-machined parts required to make up the bearing.Clearance tolerances are additive in the built-up assembly of shell,pivots, and pads, requiring a high degree of manufacturing accuracy toyield acceptable radial shaft clearances. Pad pivot friction under highradial load can also lead to premature wear, or even fatigue failure,which can enlarge clearances and increase rotordynamic unbalanceresponse. All of these requirements combine to make the tilt-pad bearingone which demands maximum attention to design, manufacturing, andmaterials.

Many of today's modern turbomachines, especially those running at highspeeds and low bearing loads, require the superior stabilitycharacteristics of tilt-pad journal bearings to prevent rotordynamicinstabilities. Until now, the design complexity of tilt-pad bearings hasprecluded their use in many small, high-volume applications where costand size are important.

The present inventor has developed an improved, less complicated movingpad bearing construction. For example, U.S. Pat. No. 4,496,251 a padwhich deflects with web-like ligaments so that a wedge shaped film oflubricant is formed between the relatively moving parts.

U.S. Pat. No. 4,515,486 discloses hydrodynamic thrust and journalbearings comprising a number of bearing pads, each having a face memberand a support member that are separated and bonded together by anelastomeric material.

U.S. Pat. No. 4,526,482 discloses hydrodynamic bearings which areprimarily intended for process lubricated applications, i.e., thebearing is designed to work in the working fluid. The hydrodynamicbearings are formed with a central section of the load carrying surfacethat is more compliant than the remainder of the bearings such that theywill deflect under load and form a pressure pocket of fluid to carryhigh loads.

It has also been noted, in Ide U.S. Pat. No. 4,676,668, that bearingpads may be spaced from the support member by at least one leg whichprovides flexibility in three directions. To provide flexibility in theplane of motion, the legs are angled inward to form a conical shape withthe apex of the cone or point of intersection in front of the padsurface. Each leg has a section modulus that is relatively small in thedirection of desired motion to permit compensation for misalignments.

U.S. Pat. No. 5,054,938 also to Ide discloses a number of bearingsparticularly well-suited for high speed equipment. The bearings includefluid dampened support structures.

Such deflection pad bearings offer exceptional damping characteristics.In addition to the damping typically associated with tilt pad bearings,the support structure and fluid located between the webs also providedamping. It is even possible to provide an oil filled diameter membraneto increase damping. Moreover, because these bearings function withoutcontact between moving parts they offer the possibility of virtuallyinfinite life.

Despite the advantages offered by these bearing constructions, they havenot yet been universally accepted. This can be attributed, at least inpart, to the revolutionary nature of these bearings and the fact thatthey are a radical departure from "conventional" thinking in the fieldof rotordynamics. In addition, when a fluid film bearing fails it oftencan completely seize without warning. The results could be catastrophicin a jet engine, for example. On the other hand, failure of a rollingelement bearing is usually gradual and indicated by the increasing noisegenerated by the bearing. Moreover, rolling element bearings work, tosome extent, even without lubricant. This certainly accounts for thecontinued use of rolling element bearings in jet engines, but does notexplain the continued use of rolling element bearings in applicationswhere loss of lubricant is less catastrophic, e.g., aeroderivativeapplications. For whatever reason, there remains a preference among somein the field for rolling element bearings. There is a need, therefore,for a simple inexpensive reliable system which provides good dampingcharacteristics for both rolling element bearings and tilt pad bearings.

SUMMARY OF THE INVENTION

The present invention discloses a fluid dampened support primarilyintended for a rolling element bearing, but which can also be used inconjunction with an advanced tilt pad bearing construction and methodsof making the same. The bearings offer many of the rotordynamicadvantages of tilt-pad bearings without sacrificing the advantages ofrolling element bearings. This is achieved by applying the deflectingsupport-principle, previously used on movable pad bearings to rollingelement bearings. These principles are described in co-pending parentapplication Ser. No. 07/309,081 filed Feb. 8, 1989 now U.S. Pat. No.5,137,373 the complete disclosure of which is incorporated herein byreference. Thus, the support member which is preferably unitary, can beformed from a single piece of heavy walled tubing or a cylindricaljournal that has been machined or formed with small grooves and slits,bores or cuts through or on the bearing wall to define a support surfaceand a flexible support structure.

All of the essential parts of the support structure are integrallyformed from a single piece of material. This allows the bearing to bemanufactured for much less than conventional designs. Moreover, becauseall of the component parts are integral, their position relative to eachother is fixed. This allows the bearings to be manufactured to muchcloser tolerances than multipart assemblies where the tolerances of eachpart are additive.

The present invention also relates to a dampening structure into which arolling element or tilt pad bearing can be inserted. Virtually anyreasonable dampening characteristic can be developed by controlling thestructure and space between beams. Additionally, a flexible fluiddampened membrane may be used to support the support structure at theouter diameter.

The fluid dampened rolling element bearing according to the presentinvention includes an inner race; an outer race; and a series of ballsor other rolling elements disposed between the inner race and the outerrace such that the inner race is rotatably supported on the outer raceby the rolling elements. A support structure is provided for supportingthe outer race for radial and torsional movement. The support structureperforms the function of a squeeze film centering ring in conventionaldamper assemblies. The support structure comprises a plurality ofcircumferentially spaced members separated from one another by narrowspaces. An incompressible fluid is provided in these spaces to allowfluid dampening.

Likewise, the support structure can be used to support a one-piece tiltpad bearing which includes a plurality of substantially circumferentialbearing pads and a bearing housing extending radially outside thebearing pads and encircling the bearing pads with a plurality of thinwebs equal in number to the number of bearing pads, each web extendingradially between one of the pads and the bearing housing so as tosupport the pad on the bearing housing for pivoting movement. The outerhousing is supported by the dampening support structure mentioned above.

In accordance with an important aspect of the present invention, thedamper for supporting either a rolling element bearing or a one-piecetilt pad bearing includes a plurality of support pads for supporting thebearing. Each of the support pads include a radially inner surfacesupporting the bearing and a radially outer surface and spacedcircumferential ends. A structural spring is provided to support each ofthe circumferential ends of each of the bearing pads so as to allowradial movement of the pads and, if desired, torsional movement of thepads. Each of the structural springs has a predetermined springconstant. The present invention provides a means for changing the springconstant of the structural spring after the spring has deflected apredetermined amount. Specifically, an adjustable post is provided tocontact the structurals after a predetermined amount of deflection. Thepost contacts the spring, the effective length of the spring isshortened so that the spring constant is increased. By adjusting thespace between the post and the spring the amount of deflection thatoccurs before the onset of the change in spring constant and be varied.Accordingly, the post is preferably mounted in an adjustment means suchas threaded assembly or the like to allow movement of the post towardand away from the structural spring.

The damper of the present invention further includes a generallycircular base supporting each of the structural springs. The base has aradially inner surface and a radially outer surface. The radially outersurface of each of the pads is spaced from the radially inner surface ofthe base so as to define a circumferentially extending damping gap. Thedamping gap is filled with an incompressible fluid so as to dampenmovement of the pad with respect to the base. An orifice is provided atthe circumferential ends of the gap to allow the incompressible fluid toflow into and out of the gap in response to movement of the pad relativeto the base. The orifice is preferably defined by radially inwardextending projections formed on the radially inner surface of the base.The size of the orifice is determined by the space between thecircumferential edge of the pad and the surface of the projection.According to the present invention, the shapes of the circumferentialedge of the pad and the projection can be designed such that the size ofthe orifice varies with deflection of the pad toward or away from thebase. Specifically, the radially inward extending projection can betapered away from the pad such that the orifice becomes smaller with paddisplacement. Alternatively, the projection can be tapered toward thepad so that the orifice increases with pad displacement. Finally, theprojection and edge can be non-tapered so that the size of the orificeremains constant with pad deflection.

The assembly is designed such that a space is provided between everypoint on the outer race and the outer periphery of the support member.If desired, the support structure may be in the form of a one-piecemember formed with cuts and grooves to define a plurality of beams. Thecuts and grooves can be formed through electric discharge machining soas to provide very narrow openings to enhance the fluid dampeningeffect. The support structure can be formed separate from the outer raceto allow fluid dampened support of conventional, off the shelf rollingelement bearings. Moreover, the support structure can include acontinuous inner ring for supporting the outer race of the rollingelement bearing or alternatively the inner ring could be formed by aseries of special pads rather than a continuous ring. Otherconstructions are, however, possible. For example, the continuous innerring and the outer race of the bearing could be integrally formed.

The support structure may include a fluid dampened membrane so that thesupport structure functions as a squeeze film damper. The supportstructure is designed to optimize the damping characteristics of thesupport structure. This can be done by modifying the support structure.The support structure can be designed to support the support surface formovement in up to six degrees of freedom (i.e., translation or movementin the +x, -x, +y, -y, +z and -z directions) and rotation about the X,Y, and Z axes so as to ensure damping at all times.

The support structure is preferably unitary (one-piece) and comprisessupport stubs, beams, and/or membranes connected to a housing which issometimes defined by the radially outermost portion of the bearing inthe case of a journal bearing or, in the case of thrust bearings, ahousing into which the bearing is mounted.

The inventor has discovered that in many specific applications such asin high speed applications, it is necessary to examine and evaluate thedynamic flexibility of the entire system consisting of the shaft orrotor, the bearing and the fluid dampened support structure. In computeranalysis of this system using a finite element model, it has beendetermined that it is necessary to treat the entire support structure asa completely flexible member that changes shape under operating loads.By adding more or less flexibility via machining of the basic structure,damping characteristics may be achieved that provide stable low frictionoperation over wide operating ranges. A number of variables have beenfound to substantially affect the support structures dampingcharacteristics. Among the most important variables are the shape, size,location and material characteristics (e.g. modulus of elasticity etc.)of the support members defined by the bores, slits or cuts and groovesformed in a one piece member. The shape of the support members has beenfound to be particularly important. Also by providing a fluid backing tothe flexible members, a high degree of damping may be achieved thatfurther adds to system stability.

While there are numerous arrangements of bores, grooves, cuts, or slitsthere are primarily two modes of deflections: namely, one or moreligaments or membranes which deflect in the general direction of load ina bending mode and secondly, by torsional deflection in a beam ormembrane in a direction extending away from the support surface alongthe longitudinal axis of the shaft. The degree of deflection in thebending mode is, in part, a function of the stiffness of the supportstructure in the radial direction. The cuts are specifically made toresult in a predetermined shape under load. By surrounding or backingcertain ligaments or membranes with lubricating fluid, an additionaldamping element may be added to the design.

Though various designs are possible, there are two key performancecharacteristics which the damper should have. First, the structureshould have sufficient flexibility to ensure that actual squeezing ofthe fluid film occurs. Second, the damper should be capable of dampingunbalanced loads in all directions.

The thickness of the gap between damper parts (corresponding to thesqueeze film thickness) and the required flexibility are necessarilyinterrelated. The clearances between the damper parts must be smallenough that the flexibility allowed by the support structure issufficient. It is noted, in this regard, that pressure is a cubicfunction of thickness, i.e., pressure is proportional to the thirdexponential power of thickness. If the gap is very small then thesupport structure does not have to be so flexible.

Because the support structures of the present invention are essentiallyone-piece continuous beam networks, they are, of course, relativelyrigid in comparison to soft spring dampers. This rigidity offerssignificant advantages in terms of stability and wear. On the otherhand, it requires a relatively small space between the damper parts.Thus, to achieve proper squeeze film characteristics with the bearingsof the present invention, the largest allowable space between damperparts is normally 3-5 mil and certainly not more than 8 mil. This doesnot present a significant problem because, in accordance with thepresent invention, the cuts which define the space between damper partscan be made using wire cut EDM (electrical discharge machining). ManyEDM machines are capable of making cuts as small as 1 or 2 mil. If othermachining techniques are used, such as a conventional wire cut EDM whichleaves a gap of 10 mil or more, the space can be reduced by placing aseparate shim in the gap.

In the support structures of the present invention, the flexibilitydepends primarily on the length and cross-section of the beams orligaments which support the damper parts. The deflection characteristicsof any particular beam configuration can easily be obtained from astructural engineering handbook.

The bearing assembly of the present invention is especially well-suitedfor use in jet engines where damping is critical. Other specificapplications of the bearings of the present invention include electricmotors, fans, turbochargers, internal combustion engines, outboardmotors, and compressors/expanders.

The support structure member may be formed of metals, powdered metals,plastics, ceramics or composites. The bearing can be tuned by, forexample, changing the support structure and gaps to alter the stiffness.This in turn eliminates vibration. The present invention alsocontemplates easily moldable support structures which include no hiddenopenings such that they can be molded in a simple two-piece mold.

Various methods of manufacturing the supports of the present inventionare also contemplated. The selection of a particular method ofmanufacturing depends largely on the volume of the particular support tobe manufactured and the materials used.

In low volume applications, or when it is desired to produce prototypesfor testing and/or production of molds or the like, the supports arepreferably manufactured from metallic cylindrical blanks such as heavywall tubing or other journals which are machined to provided radialand/or facing bores or grooves and formed with radial cuts or slitsthrough either numerically controlled electrical discharge manufacturingtechniques, numerically controlled laser cutting techniques, ornumerically controlled water-jet cutting.

In intermediate volumes, the supports of the present invention arepreferably manufactured using an investment casting method in accordancewith the present invention.

In high volume applications, the supports of the present invention canbe manufactured using a wide variety of materials such as plastics,ceramics, powdered and non-powdered metals, and composites. In highvolume applications, a number of manufacturing methods, includinginjection molding, casting, powdered metal, die casting, and extrusion,can be economically employed. The supports of the present invention canbe formed in a shape which is easily moldable.

The present invention offers a number of advantages over conventionaldesigns. Unlike squirrel cage spring dampers, the present invention doesnot require any additional axial space beyond that already provided bythe bearing. Unlike elastomer O-ring dampers, the present invention canaccommodate a very wide range of stiffness values which can be predictedaccurately and will not change with time or temperature. The damper ofthe present invention is also capable of taking a thrust load. Moreover,unlike many of the designs used with fluid film bearings which utilize amultiple of sector beams to provide a flexural support, the presentinvention includes a one piece support structure which can also be splitfor ease of assembly. The one piece construction allows for greaterprecision and ease in assembly by avoiding the problems of tolerancestack up which are inevitable with multipart designs. Further, unlikeany of the existing squeeze film damper bearing designs, the presentinvention allows for changing of the squeeze film clearance space by,for example, inserting shims in the space to fine tune the design.

In operation, the construction of the present invention can isolate thesqueeze film region into several pockets and prevents interactionbetween the cavities. This allows for maximizing of the damping that canbe obtained from the damper. In the design of the present invention, thestatic weight offset can be accurately accounted for during themanufacturing of the ring. No special assembly routines will be neededto account for the static weight offset. The support structure canreadily accept any existing antifriction bearing or fluid film bearingwith only minor modifications to the bearing housing.

Finally, unlike any of the existing squeeze film damper designs, thepresent invention allows for only positive pressure generation in thedamper bearing. The squeeze film cavities are designed so thatseparation of the surfaces does not follow as the journal moves away.This prevents the formation of negative pressure in the squeeze filmcavity. Therefore, cavitation and air ingestion, which is a primaryfactor for the poor performance in most damper bearings, is completelyeliminated.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a sectional view of a rolling element bearing assembly whichincludes one form of fluid dampened support structure according to thepresent invention;

FIG. 1A is a sectional view along the lines indicated in FIG. 1;

FIG. 2 is a sectional view of a second rolling element bearing assemblyaccording to the present invention;

FIG. 2A is a sectional view along the lines indicated in FIG. 2;

FIG. 3 is a side view of another fluid dampened support structureaccording to the present invention;

FIG. 3A is a radial cross-section of a portion of the support structureillustrated in FIG. 3;

FIG. 4 is a sectional view of a rolling element bearing having a fluiddampened support structure integrally formed with the outer race;

FIG. 4A is a sectional view along the lines indicated in FIG, 4;

FIGS. 5A and 5B are cross sectional views of a cylindrical journal orblank prior to machining;

FIGS. 6A and 6B are cross sectional views of a machined journal orblank;

FIGS. 7A and 7B are cross-sectional views of a further machined journalor blank;

FIGS. 8A and 8B are cross sectional views of a modified machined journalor blank;

FIGS. 8C and 8D are cross sectional views of a support structureconstructed from the modified machined journal or blank of FIGS. 8A and8B;

FIG. 9 is a sectional view of another rolling element bearing assemblyaccording to the present invention;

FIG. 9A is a detail view showing one section of the support structure ofthe assembly shown in FIG. 9;

FIG. 9B is a simplified model of the damper structure depicted in FIG.9A; and

FIG. 9C is a schematic representation of the damper structure depictedin FIG. 9B.

FIG. 9D is a sectional view of a fluid dampened one-piece tilt padbearing according to the present invention.

FIG. 10 is a sectional view of a damper structure with a variablestiffness characteristic.

FIG. 10A is a detail view of the damper structure of FIG. 10.

FIG. 10B is a graph showing force as a function of displacement toillustrate various spring characteristics.

FIG. 11 is a sectional view of a damper structure with a variableorifice.

FIG. 11A is a detail view of another damper structure with a variableorifice.

FIG. 11B is a detail view of another damper structure with a variableorifice.

FIG. 11C is a detail view of a damper structure with a constant orifice.

FIG. 12 is a sectional view of a damper structure having both adjustablestiffness and a variable orifice.

FIG. 12A is a detail view of the damper of FIG. 12.

DETAILED DESCRIPTION OF THE ILLUSTRATIVE EMBODIMENTS

In describing the bearings and especially the support structure of thepresent invention in an understandable way, it is helpful to describethe support structures as being formed from a cylindrical blank byproviding grooves, slits, bores and other openings in the cylindricalblank. As noted below, this is sometimes a useful technique formanufacturing a prototype support structure. The reference to thecylindrical blank is, however, primarily intended to assistunderstanding of the present invention. Although many of the supportstructures of the present invention could be manufactured from acylindrical blank, it is not necessary that any of them be somanufactured. Indeed the support structure can be manufactured innumerous ways, some of which are discussed hereinafter.

Thus, the support structure may be described as a journal machined todefine an inner peripheral ring or set of pads for supporting the outerrace of the rolling element bearing, an outer periphery supported inhousing and a network of beams and membranes providing flexible supportfor the inner periphery, and hence the bearing relative to the outerperiphery and housing.

The primary flexibility of the support structure of the presentinvention is developed by small cuts or slits through the journal wall.These cuts provide the inner peripheral ring or pads with up to sixdegrees of freedom (i.e., the ring or pads can translate in the +x,-x,+y,-y, +z and -z directions as well rotate about the x, y and z axes)and are designed to optimize the damping characteristics of the supportstructure. The cuts or slits can be provided to provide a continuouscylindrical membrane. The membrane acts as a fluid damper upon which theremainder of the support structure and the rolling element bearing aresupported. The flexibility of the membrane, combined with the fluidlubricant, provides a means to vary the damping action and to isolatethe pad from the housing. The damping takes the form of a dash pot thatexhibits high damping characteristics.

Referring first to FIG. 1, one of the currently preferred embodiments ofthe present invention is shown. The roller bearing assembly showntherein includes conventional rolling element bearing components such asthe inner race 11, outer race 12 and rolling elements 13 such as ballsor rollers. These components are of a conventional construction. Indeed,this portion of the assembly may be provided through the use of a simple"off the shelf" ball, roller, or needle bearing assembly. Many suchassemblies include additional components such as, for example, a cage toseparate the rolling elements from one another.

In accordance with the present invention, the conventional rollingelement bearing components 11, 12 and 13 are supported by fluid dampenedsupport structure generally indicated at 20. The support structure 20 isformed with grooves and slits so as to define a continuous inner ring21, an outer periphery 23 and a network of circumferentially spacedbeams 24, 26 and 28. the inner ring 21 serves as a support surface forsupporting the outer race 12. The outer periphery 23 is adapted to besupported in a rigid housing or the like. The network ofcircumferentially spaced beams 24, 26 and 28 support the radially innersupport surface 21 and hence the inner race 11, outer race 12 androlling elements 13 of the assembly for movement relative to the outersurface 23.

As shown in FIG. 1, the beam network includes a series of eightstub-like beams 24 which provide the only connection between the beamnetwork and the inner support ring 21. Each of the stub-like beams 24 isconnected at one end to the inner ring 21 and at its other end to acircumferentially extending beam 26. The circumferential beams 26 areeach connected to the stub section or beam 24 at one end and to a secondstub-like beam 28 at the opposite end such that the beams act ascantilever supports. The beam sections 28 are in turn connected to thecircumferential beams 26 at one end and to the outer periphery 23 attheir opposite end. As can be seen from FIG. 1, the cuts and slits arearranged to provide an open space between every point on the innersupport surface and the outer periphery of the support structure. Thisgives the support structure significant radial flexibility.

If additional flexibility is desired, facing grooves can be formed onthe sides of the support structure to reduce the torsional stiffness ofthe support structure. In particular, as shown in FIG. 1A axiallyextending facing grooves can be provided on each side of the supportstructure so as to reduce the axial dimension of one or more of thebeams and the support structure thereby reducing the torsional stiffnessof the beam.

If even more dampening is desired, a radial extending groove can beformed in the outer periphery of the support structure 23 so as todefine a membrane upon which the beams are supported. Specifically, asshown in FIG. 1A the provision of a groove results in formation of twocircumferential beams 23a, 23b which are connected to another body andmembrane the thickness of which is dependent upon the depth of thegroove.

When used in high speed equipment such as jet engines, the rollingelement bearings are typically located in a sealed chamber which isflooded with lubricant to remove the heat generated by the movement ofthe rolling elements. When the rolling element bearing assembly of thepresent invention is located in such a flooded chamber the lubricantwill naturally fill the spaces between the various beams. The fluidlocated in the interstices of the support acts like fluid in a dashpotto provide dampening of the movement of the beams in the supportstructure.

Another currently preferred embodiment of the support structure of thepresent invention is shown in FIG. 2. As shown there, the rollingelement bearing assembly, illustrated somewhat schematically, againincludes an inner race 11 and outer race 12 and rolling elements 13. Thesupport structure generally indicated at 120 includes a continuous innerring 121, an outer periphery 123 and a network of beams 124, 126 and128. Again, the network of beams is formed by a spaced series of slitsformed in a continuous journal. In this case, however, the slits are allcircumferential as shown in FIG. 2. The nature of these supportstructures is, however, essentially the same. Specifically, the supportstructure includes a series of, in this case, four stub sections orbeams 124. A series of circumferential beams 126 connected at one end tothe stubs 124 and at the opposite end to stubs 128. The stubs 128 areconnected at their opposite end to the outer periphery 123. From FIG. 2,it can be seen that each stub 124 has two circumferential beams 126extending therefrom and that each of the secondary stubs 128 supportstwo circumferential beams 126. Again, the support structure is arrangedso that there is an open space between every point on the inner ring 121and every point on the outer periphery 123.

Additionally, as with the previous embodiment, if desired, torsionalflexibility can be added to the support structure by providing axiallyextending facing grooves on the sides of the support structure.Specifically, as shown in FIG. 2A the support structure can be providedwith facing grooves to reduce the torsional stiffness of one or more ofthe beam elements. If even more damping is desired, a radially inwardextending groove may be formed in the outer periphery 123 so as toprovide a membrane support for the beam network. The provision of theradially inwardly extending groove defines two circumferentiallyextending beams 123a, 123b supporting the membrane as shown in FIG. 2A.

When the bearing assembly shown in FIG. 2 is located in a floodedcontainer such as that used in high speed applications, the lubricantwill fill the gaps between the beam elements and also fill the baseunderlying the membrane. In this way, the lubricant will act as a fluiddash pot to dampen movement of the support structure elements.

The previous embodiments disclose support structures which areparticularly well suited for retrofit applications in which theconventional bearing is provided with a separate support structure toprovide the desired damping. This is currently the most likelyapplication of the invention. It should be understood, however, that theinner ring 21, 121 of the support structures disclosed herein could, ifdesired, be used as the outer race of the rolling element bearingassembly. An example of such a construction is shown in FIGS. 4 and 4Aand discussed below. Such an assembly can simplify the assembly ifassembly were to be built entirely in one location. Of course, the outerrace of typical rolling element bearings must have certain materialcharacteristics which might require more expensive material than onemight want to use on the support structure. Accordingly, it still may beadvantageous to form the outer race and the inner ring of the supportstructure separately.

In addition, in the previously described embodiments, the inner supportring is continuous. This ensures that the rolling element bearing andparticularly the outer race can be securely retained by the supportstructure. It is, however, possible to use support structures in whichthe outer race is supported in a circumferentially spaced pads ratherthan on a continuous ring. Support structures having such acircumferentially spaced series of pads can be designed along theprinciples described in the previous application and used for bearings.

The use of separate pads rather than a continuous ring offers bothadvantages and disadvantages. One disadvantage associated with supportstructures having separate pads for supporting the outer race is that asthe deflection of one pad occurs, the tightness of the fit of anotherpad is loosened. To accommodate this, the outer race must be pressfit tosome extent in the support structure. There is also a greaterpossibility for unbalanced damping when separate pads are used. Thiscan, however, be minimized by designing the support structure such thatthe loads are evenly shared. A major advantage of separate pad supportsis that their performance is more easily modeled and reliably predicted.Additionally, separate pad supports have greater flexibility. At presentboth continuous and separate pad supports appear to be promising.Accordingly, there is no universal preference for one or the other atthis time. The selection of one or the other must therefore be made on acase by case basis taking the foregoing into consideration.

FIGS. 3 and 3A illustrate one such separate pad embodiment. Morespecifically, FIGS. 3 and 3A illustrate the possibility of using ajournal bearing as multipad support structure in accordance with thepresent invention. It should be appreciated, however, that a widevariety of support structure arrangements with a separate pad could beconstructed in accordance with the principles previously set forth inthe application incorporated herein by reference.

The construction illustrated in FIGS. 3 and 3A is bi-directional, i.e.,the structure is symmetrical about its center line. Like the previouslydescribed supports, the support of FIGS. 3 and 3A is formed with aplurality of thin radial and circumferential slits. In this case,however, the slits define a plurality of circumferentially spacedsupport pads 32.

The structure supporting each of the support pads 32 is such that eachpad 32 is supported by a beam support structure at two pad supportsurfaces 32ps. The beam network connected to the support pads at eachpad support surface 32ps is identical, yielding the symmetricalconstruction which makes the support bi-directional. For purposes ofsimplifying this description, only the network of beams which supportsthe pad at one pad support surface will be described since the other padsupport surface supported in an identical fashion. Thus, as shown inFIGS. 3 and 3A, a first, generally radially extending beam 40 isconnected to the pad 32 at the pad support surface 32ps. A second,generally circumferential beam 42 is connected to the radially outermostend of beam 40. A third, generally radial, beam 44 extends radiallyinward from the beam 42. A fourth, generally circumferential beam 46extends from the radially innermost portion of the beam 44. A fifth,generally radial beam 48 extends radially outwardly from a beam 44 tothe housing portion 47 of the support structure. In summary, each pad 32is supported by ten beams and the bearing housing.

Further, by forming radially extending circumferentially spaced groovesor continuously extending circumferential grooves in the housing portionof the support structure, the housing portion of the support structurecan be designed to act as a plurality of beams or membranes. Thus, as aresult of the beam on beam two point support, the pad acts like aspring-like membrane.

As mentioned above, in some instances it may be desirable to make theinner support ring of the support structure integral with the outer raceof the rolling element bearing. FIGS. 4 and 4A illustrate such anassembly. As shown therein, assembly is essentially identical to that ofFIGS. 2 and 2A except that the outer race of the bearing and the innerring of the support structure are a single piece.

The assembly shown in FIG. 4 is not, however, currently preferred. Asknown to those skilled in the bearing art, the races of rolling elementbearings must be manufactured to exacting specifications and often madeof durable materials. It is currently expected that providing such arace integrally with the support structures of the present inventionmight be unduly complicated. For these reasons, the integralconstruction shown in FIGS. 4 and 4A is not currently preferred.Nevertheless, it should be appreciated that construction shown in FIGS.4 and 4A operates in essentially the same way as the construction shownin FIGS. 2 and 2A.

While certain examples are described above, it should be appreciatedthat numerous modifications to the support structure are possible. Forexample, the deflection and damping characteristics of the supportstructure can be modified by changing the angle of the beams, changingthe location of the holes or openings which define the legs, varying thelength of any of the beams or membranes, and changing the width orthickness of any of the beams or membranes. Other possible modificationsof beam mounted support structures are described in co-pendingapplications Ser. No. 309,081 incorporated by reference above.

As noted earlier, there are two key performance characteristics whichthe damper should have. First, the structure should have sufficientflexibility to ensure that actual squeezing of the fluid film occurs.Second, the damper should be capable of damping unbalanced loads in alldirections.

The thickness of the gap between damper parts (corresponding to thesqueeze film thickness) and the required flexibility are necessarilyinterrelated. The clearances between the damper parts must be smallenough that the flexibility allowed by the support structure issufficient.

Because damping pressure is a cubic function of thickness, the gap mustbe very small to accommodate the moderate flexibility of the supportstructures of the present invention. To achieve proper squeeze filmcharacteristics with the bearings of the present invention, the largestallowable space between damper parts is normally 3-5 mil and no morethan 8 mil. Accordingly, the cuts which define the space between damperparts are preferably made using wire cut EDM (electrical dischargemachining). Many EDM machines are capable of making cuts as small as 1or 2 mil. If other machining techniques are used, such as a conventionalwire cut EDM which leaves a gap of 10 mil or more, the space can bereduced by placing a separate shim in the gap.

In the support structures of the present invention, the flexibilitydepends primarily on the length and cross-section of the beams orligaments which support the damper parts. The deflection characteristicsof any particular beam configuration can easily be obtained from astructural engineering handbook.

Taking these factors into account another, currently preferred,embodiment of the present invention which is depicted in FIGS. 9-9C willbe described. As shown in FIG. 9, this embodiment is of a pad typesupport structure. The support structure includes four circumferentiallyspaced pads 232. Each of the pads 232 rest on a pad surface 232ps whichis supported by a dog legged shaped beam network which acts as astructural spring. More specifically, the beam network includes acircumferential beam 234 extending from each circumferential end of thepad support surface 232ps and a radial beam 236 extending radially at asharp angle away from the circumferential beam 234 to the outerperiphery portion 220.

As can be appreciated from FIGS. 9 and 9A, the beams 234 and 236 arequite thin. On the other hand, the beams are relatively short. Thinbeams tend to be more flexible, but short beams tend to be lessflexible. Thus, the support structure as a whole would have a moderateflexibility. The specific dimensions necessary can be determined eitherthrough trial and error or, preferably, through finite element analysis.The pad member 232 also includes a damper portion 232d at the radiallyoutermost portion thereof. The damper portion 232d extends radiallyoutward to define a thin squeeze film gap with the outer peripheryportion 220. As mentioned before, the gap should be, normally, in therange of 3-5 mil. The gap is filled with a hydraulic fluid 70 or thelike.

The support structure is essentially formed from a single piece. If itis desired to control deflection of the radial beams 236 in onedirection, however, deflection control inserts 240 can be inserted intothe support structure to prevent the beam 236 from deflecting away fromthe beams 234. As explained below, such deflection is neither necessarynor desired to achieve damping performance. Instead, beam 236 mustdeflect in the direction indicated in FIG. 9B to achieve proper damping.

As shown in FIG. 9, the pads 232 of the support structure are formedwith a larger diameter than the outer diameter of the rolling elementbearing to provide a positive preload. Such machining of differentdiameters for each of the pad is somewhat more expensive than simplymilling a uniform diameter, but such pad construction offers aperformance advantage in the construction shown in FIG. 9. Inparticular, loads from the shaft or bearing are received on or near thecircumferential center of the pad 232 so that the pad acts as a damperand deflects virtually radially rather than with a wedge shape as in ahydrodynamic bearing. In other words, supporting the outer race only ator near the center of the pads causes the force on the support structureto act at or near the center so that the structure functions as shown inFIGS. 9B and 9C with the load applied centrally. This ensures properdamper performance.

In the structure shown in FIGS. 9 and 9A, the support structure, i.e.,squeeze film centering spring, includes four distinct segments whichform a continuous ring. Naturally, the number of segments can be variedif desired. It should be noted, however, that damping values can becomeundesirably low if too many segments are used.

With the assembly shown in FIG. 9 and 9A, as the bearing whirls orvibrates, it tends to displace one or two of the segments at anyparticular time. This will cause the segments to move radially andsqueeze the oil 70 in the small clearance cavity. The squeezing actionwill generate a pressure which when multiplied by the area results in aforce proportional to the velocity of the journal. This force is adamping force that tends to dampen and reduce the vibration levels inaddition to the forces transmitted to the bearing housing and structure.

The thickness and length of the web sections at both ends of eachsegment determines the stiffness of the support. A wide range ofstiffness values can thus be achieved with dimensional change to thesecritical sections. The squeeze film cavity can also be shimmed to varythe clearance between the damper portion and the outer periphery portionso as to fine tune the damper performance. For horizontal applications,i.e., applications in which the shaft is not displaced from thehorizontal position, the bottom two segments can be sized to account forthe static weight offset thus eliminating the need for delicate anddifficult field adjustments. Multiple cavities or segments can bedesigned so as to work in series or parallel with the cavity shown andthus provide a wider range of design capability.

In simple terms, the damper construction shown in FIG. 9 and 9A operatesas a simple fluid dashpot as represented schematically in FIG. 9C. Thiscan best be understood by reference to FIG. 9B which is a simplifiedversion of the segment depicted in FIG. 9A. Thus, although the segmentdepicted in FIG. 9A is circumferential, it is useful to consider it as astraight structure of the type shown in FIG. 9B. Because the pads 232have a larger diameter than the outer race 12 of the rolling elementbearing, the force applied to the support structure of the shaft orbearing acts at or near the circumferential center of the pad as shownin FIG. 9B. This force F causes the beams 234 and 236 to deflect in thedirection indicated by the small arrows such that the damper portion232d of the pad 232 squeezes the narrow film between it and the outerperiphery 220.

This system may be schematically represented as shown in FIG. 9C as adashpot with springs at each end of the supported member 232. In thiscase, the beams 234, 236 provide the spring function and the damperportion 232d and outer periphery 220 along with the fluid filmtherebetween provide the dashpot performance.

The damper construction of the present invention can also be used tosupport a one-piece tilt pad bearing of the type disclosed in co-pendingapplication Ser. No. 07/878,601 filed May 5, 1992.

FIG. 9D shows a one-piece tilt pad bearing supported in a damper of thetype illustrated in FIGS. 9-9C. The difference between the overallbearing assembly shown in FIG. 9D and that shown in FIGS. 9-9C residesonly in the use of a one-piece tilt pad bearing instead of a rollingelement bearing. Thus, the damper again includes four circumferentiallyspaced pads 232. Each of the pads is supported by a dog-legged shapedbeam network which acts as a structural spring. The beam networkincludes a circumferential beam 234 extending from a circumferential endof the pad and a radial beam 236 extending radially at an angle awayfrom the circumferential beam 234 to support that beam on the outerperiphery portion or base 220.

The one-piece tilt pad bearing 10 is of the type which includes aplurality of circumferentially spaced pads each of which is supported ona continuous base by a support structure which, can be in the form of amultibeam support structure or, more simply, in the form of a web orligament which is thin enough to exhibit tilt pad performance. Thenumber of pads and dimensions of the beams or ligaments can varyaccording to the needs of any particular application. In thoseconstructions in which the pads are supported on a single thin web theperformance of expensive multi-part rocking pivot pad bearings can bereplicated in a single-piece bearing. The pivot stiffness of the pads isdetermined by the support web thickness. When the web thickness is lowenough, tilt pad behavior results, i.e., the pad tilts with almost norotational stiffness.

In accordance with one aspect of the present invention, the single piecetilt pad bearing and the single piece damping support structure may beintegrally formed as a one-piece bearing and damper. Although thisconstruction is not illustrated it could be similar to the assemblyshown in FIG. 9D with the bearing 10 bearing integrally connected to thesupport pads 232 at its base 19.

FIG. 9D illustrates one example of the tilt pad bearing supported in adamper according to the present invention. The bearing 10 is designed tooperate in a liquid filled environment. In operation, the pads 15 tiltso as to pressurize the liquid. As shown, the bearing includes fourspaced bearing pads 15. The number of bearing pads can, of course, bevaried to suit any particular application. The bearing pads 15 are eachsupported on a bearing housing 19 via a single thin web-like ligament 17which extends generally radially between the pads and the bearing baseor housing. As shown, the ligament has a radial length whichsignificantly exceeds its circumferential width.

In the embodiment illustrated, the web 17 is provided on thecircumferential centerline of the pad 15 so the bearing supports theshaft for rotation in either the clock wise or counter clock wisedirection, i.e., the bearing is bidirectional. If bidirectionaloperation is not necessary, the web can be attached closer to thetrailing edge of the pad to increase wedge stiffness.

The squeeze film damper constructions shown heretofore can be modelledon a computer using finite element analysis and designed for anyparticular application. In some instances, however, it is advantageousto be able to adjust the damping characteristics after the assembly hasbeen constructed to allow fine tuning or adjustment for actual operatingconditions. Moreover, the ability to vary the stiffness of the damper'sspring makes it possible to vary the critical speed of the rotor. Thestiffness also indirectly affects the effective damping that can beobtained. A lower spring stiffness permits more damper motion and,consequently, more damping. Too much damping causes the bearing tolock-up and act like a very stiff support. Therefore, the ability tovary the damping and match the bearing damping to the particularapplication is very important. The present inventors have devisedarrangements whereby two important damper characteristics, namely springconstant and orifice size can be varied during operation.

In the squeeze film damping constructions disclosed heretofore, thedamping characteristics are set once the damper is constructed.

These arrangements are particularly well suited for a damperconstruction in which the support pads are supported by structuralsprings on their circumferential ends. The structural springs are in theform of first beam or ligament supported on another, second, beam orligament at one end and supporting the pad at its other end. The springconstant or force required per unit displacement of such a structuralspring depends on the length of the first beam.

In accordance with the present invention, the effective length of thebeam can be shortened by providing a post which is spaced apredetermined distance from the structural spring so that the structuralspring contacts the post after a certain amount of deflection. When theeffective length of the beam is shortened during deflection in this way,the spring constant is increased so that the structural spring has, ineffect, two different spring stiffnesses.

The gap or spacing between the post and the structural spring determinesthe range of radial displacement in which the softer spring constantoperates. The high stiffness range becomes active when the gap closesand the structural spring becomes shorter and stiffer. As shown in FIG.10B (discussed below), the location of the post and the gap distance canbe used to provide a wide range of stiffness characteristics in thesqueeze film damper.

The stiffening of the spring after a predetermined displacement can, forexample, be used to provide safer operation of an aircraft engine in theevent of a blade loss by preventing blade rub and reducing oreliminating the impact the rotor can make against the damper housing.

There are other advantages associated with adjustable spring stiffnessand variable damping. Spring stiffness is often the variable utilized tochange the location of the critical speed and the amount of effectivedamping in the squeeze film damper. The stiffness can also be used tocontrol the amount of rotor deflections for purposes of maintaining theblade tip clearance. Varying the amount of damping can prevent damperlock-up problems. Controlling the amount of damping can reduce theforces transmitted to the housing from the rotor. Certain damper designshave a step in the squeeze film land to prevent shaft movement beyond acertain point radially. This step will result in a non-linear springeffect and the impact of the rotor can result in undesirable rotorvibrations. The harder spring in the damper design of the presentinvention will have some non-linearity, but will have better absorptionthan the rigid step in the damper. There is no need to use a step in thedamper as a stop since it degrades the damper performance and can resultin undesirable non-linear behavior.

In addition, the present inventors have found that the orifice at theend of each pad or sector can be varied to provide different end sealconditions. This can also add stiffness to the squeeze film damper notavailable in the traditional squeeze film damper designs. The orificecan be made to become smaller, remain constant, or become larger as thejournal or pad is displaced radially. Depending on the need andapplication, the orifice configuration can be adjusted to provide theoptimum performance for the specific application. This design featurecan also be used to control the amount and extent of cavitation in thedamper since cavitation and air entrainment are influenced by theorifice size and the end seals.

Having described the general nature of the adjustability featuresdiscovered by the present inventors, specific structures incorporatingsuch features will now be described with reference to FIGS. 10-11C.

FIGS. 10 and 10A show a damper construction in which the spring rate ofthe structural support spring changes (becomes higher) at some point asthe pad moves radially. The point at which the spring rate changes canbe adjusted.

The squeeze film damper construction shown in FIGS. 10 and 10A issimilar to that shown in FIG. 9. Specifically, the damper includes fourcircumferentially spaced pads 232. Each of the pads has a radially innersurface 232i, a radially outer surface 232r and two circumferential ends232e. The pads 232 are supported on their ends 232e by a dog-leggedshaped beam network which acts as a structural spring. The beam networkincludes a circumferential beam portion 234 extending from eachcircumferential end 232e of the pad 232 and a radial beam portion 236extending radially at an angle away from the circumferential beam 234 tothe outer periphery portion or base 220.

As can be appreciated from FIGS. 10 and 10A, the beams 234 and 236 arequite thin. Thin beams tend to be more flexible, but short beams tend tobe less flexible. Thus, the flexibility of a support structure whichincludes beams of fixed thickness depends on the effective length of thebeam.

The radially outer surface 232r of each of the pads 232 is spaced fromthe inner surface of the base 220 to define a squeeze film damping gap232g. The damping gap 232g is filled with an incompressible fluid toprovide damping in a manner known in the art. As best shown in FIG. 10A,the damping gap 232g opens into a wider chamber at an orifice 232oformed at each circumferential end of the squeeze film damping gap 232g.

In accordance with the present invention, an adjustable post assembly50, best shown in FIG. 10A, is associated with each structural spring(234 and 236). The adjustable post assembly 50 includes a springcontacting post 52 and a threaded adjustment assembly 54 for adjustingthe gap between the post 52 and the structural beam 234 of thestructural spring assembly 236.

The post 52 extends toward the beam 234, but is spaced therefrom by apredetermined gap. When the pad 232 is loaded and moves toward the base220, the beam 234 deflects downward narrowing the gap between the beam234 and the post 52. As long as there is a space between the post 52 andthe beam 234, the beam 234 deflects downward at a first, lower, springrate, because the beam 234 is relatively long. When the post 52 comesinto contact with the beam 234, however, the spring rate becomes highersince the effective length of the beam 234 is substantially shortened.Thus, movement of the pads 232 toward the base 220 occurs with twodistinct spring rates. The first, lower, spring rate occurring duringthe initial displacement and being determined by the length andthickness of the beam 234. A second, higher, spring rate occurring aftercontact with the post 52 and being determined by the thickness of thebeam 234 which is unchanged and the length thereof which issignificantly shortened.

The higher spring rate is, therefore, dependent on the distance betweenthe circumferential end and the point at which the post contacts thespring. This, in turn, depends on the location of the post with respectto the spring. Thus, the location of the post assembly with respect tothe spring is selected to obtain the desired spring rate after the postcontacts the spring.

The point at which the spring rate changes from the relatively lowspring rate to the relatively high spring rate depends on the gapbetween the post 52 and the beam 234 prior to any displacement.Specifically, as the space becomes larger, a greater amount ofdeflection occurs before onset of the change in spring rate.

The post assembly 50 of the present invention includes a threadedadjustment assembly 54 which allows the post to be moved toward or awayfrom the beam 234 within certain limits so as to cause the gap betweenthe post 52 and the beam 234 to change. As a result, the point at whichthe transition between the low spring rate and the high spring rateoccurs can be adjusted to suit any particular application.

FIG. 10B graphically illustrates various deflection characteristicswhich can be achieved using the construction of the present invention.In the graph displacement is shown as a function of force. Initialdisplacement occurs at a lower spring rate. At some point along thedisplacement axis, the higher spring rate takes effect and a greaterforce is required for displacement (i.e., the spring rate is increased).

FIG. 10B illustrates four possible arrangements. In each arrangement,the lower spring rate is identical and the higher spring rate isidentical. The difference is the point at which the transition betweenthe lower spring rate and the higher spring rate occurs. Since thispoint depends on the initial gap between the post 52 and the beam 234,spring characteristic SC1 represents a situation where the post isrelatively close to the beam 234 whereas the lines indicating the springcharacteristics SC2, SC3 and SC4 correspond to a progressively greaterspace between the post 52 and the beam 234. The final line SC4corresponds to a relatively large gap between the post 52 and the post234 such that a good deal of displacement occurs before the higherspring rate takes effect.

The higher spring rate has the same slope in each case because thecircumferential point of contact of the post on the spring is unchanged.To obtain a different spring rate slope, this location must be changed.

FIG. 11 shows a squeeze film damper construction similar to that shownin FIG. 10 in which the post assembly is not shown. The embodimentillustrated in FIG. 11 further differs from that shown in FIG. 10 inthat protrusions 220p are formed on the radially inner surface of thebase 220. These protrusions are formed so as to define an orifice 230owhich controls the flow of fluid from the damping gap 232g into and outof the chamber. Any circumferential fluid flow passing between thedamping gap 232g and the outer chamber must pass through thisrestriction.

The present inventors have discovered that, in certain instances, it isadvantageous to vary the size of the orifice with displacement tothereby control flow into and out of the damping gap with displacement.FIGS. 11A and 11B show constructions which make such a variationpossible. Specifically, the protrusion 232p and the circumferential end232e of the pad 232 are configured such that the gap or orifice 232oeither increases or decreases in size as the pad moves relative to theprotrusion.

In FIG. 11A, the protrusion 232 is tapered toward the pad and thecircumferential end of the pad 232 is tapered toward the protrusion suchthat both the protrusion and the end of the pad have edges which extendat an acute angle away from one another. Thus, the passage between thedamping gap 232g and the chamber is acute and as the pad 232 movesradially outward, the gap or orifice increases with displacement.

FIG. 11B shows the opposite construction in which the protrusion 220p istapered away from the circumferential end of the pad edge 232e and thepad edge 232e is likewise tapered away from the protrusion 220p so thatthe edges of the pad and protrusion are obtuse. Thus, the passagebetween the gap and the chamber is obtuse and the gap or orifice becomessmaller as the pad 232 moves radially toward the base 220.

FIG. 11C shows an arrangement in which the edge walls of the protrusion220p and the circumferential ends of the pad 230 extend at right anglesradially and parallel such that the size of the orifice remains constantwith displacement.

FIGS. 12 and 12A show a squeeze film damper construction which includesboth a variable orifice and an adjustable spring rate. The variablespring rate construction is identical to that shown in FIG. 11B anddescribed above using the same reference numerals. The protrusion 220pis tapered away from the pad end 232e so that the orifice 232o becomessmaller as the pad 232 moves downward. The damper construction shown inFIGS. 12 and 12A also include an adjustable post assembly identical tothat shown in FIGS. 10 and 10A and described above using the samereference numerals. Thus, the advantages of an adjustable stiffnesscharacteristic and variable orifice size can be combined in a singledamper.

An important aspect of the present invention is the disclosure ofmachinable support shapes, i.e., support shapes which can be produced bymachining a piece of heavy walled tubing or similar cylindrical journalusing standardly available machining techniques. Such supports arecharacterized by the fact that they are formed from a piece of heavywalled tubing or similar cylindrical journal through the provision ofbores, slits and grooves. The advantage of such supports is that it iseasy to manufacture prototypes and to modify these prototypes aftertesting. Naturally, when the supports are to be mass produced, using,for example, molding or casting techniques, different manufacturingconsiderations may dictate different shapes. It is important torecognize that changes in shape affect support performance.

Another manufacturing consideration is ease of molding. Naturally, mostof the support structures of the present invention are capable of beingmolded by some molding technique. Only certain shapes can, however, beinjection molded in a simple two-piece mold, i.e., a mold which does notinclude cams. The supports of the present invention can be constructedwith easily moldable shapes which are defined as shapes which can beinjection molded using a simple two-piece mold. An easily moldable shapegenerally is characterized by the absence of "hidden" cavities whichrequire cams for molding. Accordingly, an easily moldable shape includesno radially extending grooves in the inner and outer diameter and acontinuous axial cross section.

The dimensions and deflection variables including number, size, shape,location and material characteristics of the elements defined in theunitary support structure can be tailored for any specific applicationto support a wide variety of loads. Of these variables, the shape of thesupport members is particularly important. The impact of shape of thesupport members on the deflection characteristics of the supportstructure can be appreciated when the variable formula for moment ofinertia bh³ /12 (English units) (the main component of sectional modulusfor rectangular section, z=I/c=bh² /6) is used as an example. Moreover,the ability of the support ring or pad to move with six degrees offreedom allows the support to compensate for and correct shaftmisalignment. In this regard it is noted that the supports of thepresent invention have a self-correcting characteristic resulting fromthe tendency of the support to return to its non-deflected state due tothe stiffness of the support. Of course, the stiffness of the support isprimarily a function of the shape of the support structure, and to alesser extent the other deflection variables, including number, size,location, and material characteristics of the elements defined by thegrooves and cuts or slits formed in the unitary element. Stiffersupports have a greater self-correcting tendency but are less able toadjust for shaft misalignment.

In small quantities, the support structures disclosed herein arepreferably constructed by electrical discharge machining or lasercutting methods. The double lines shown in the drawings are the actualpaths of the wire or beam which is typically 0.002-0.060"(0.50-1.52 mm)in diameter. The lubricant that flows into the electrical dischargemachined paths acts as a fluid dampener that reduces any vibration orinstability at resonant frequencies. In the situations described abovewhere a continuous cylindrical membrane is formed, the damping takes theform of a dash pot that exhibits high damping characteristics. With thenovel approach of tuning or modifying the stiffness of the bearingconfiguration or structure and particularly the beam to suit aparticular bearing application, optimum performance is readily obtained.Recent computer analysis has demonstrated that virtually any stiffnessor deflection may be accomplished.

As noted above, when manufacturing low volumes or prototypes of thesupport structure of the present invention, the support structure arepreferably constructed by electrical discharge machining or lasercutting methods. Such small volumes or prototypes are usuallyconstructed of metal. However, when higher volume production of aparticular bearing is contemplated, other methods of manufacture such asinjection molding, casting, powdered metal die casting and extrusion aremore economical. In connection with such manufacturing methods, it maybe more economical to use plastics, ceramics, powdered metals orcomposites to form the support structure of the present invention.Methods such as injection molding, casting, powdered metal die castingwith sintering and extrusion are sufficiently well known that theprocesses need not be detailed herein. Once a prototype bearing isconstructed, the method of producing a mold or the like for massproduction of the support structure is well known to those skilled inthe molding and casting art. Moreover, it is to be understood that onlycertain types of the support structure of the present invention areadapted to be made in high volumes through extrusion. Generally, theseare the support structures that are formed only through the provision ofcircumferential grooves and radial and circumferential cuts or slitswhich extend axially throughout the entire support structure, i.e.,those support structure having a constant or otherwise extrudablecross-section.

Investment casting may be used in the manufacture of intermediatequantities, e.g., less than 5,000 support structures. The first step ofthe investment casting procedure is manufacture of a prototype. Theprototype can be manufactured in any number of ways, but is preferablymanufactured by machining a piece of heavy walled tubing or similarcylindrical journal. In larger support structures, the cylindricaljournal typically is machined using a lathe for forming face andcircumferential grooves, and a mill for forming axial and radial bores.In machining smaller cylindrical journals, techniques such as water-jetcutting, laser and wire electrical discharge techniques are generallymore suitable. In either application, the journals are typically turnedand milled to form the larger grooves.

After the prototype is formed, it may be desirable to test the prototypeto confirm that the support structure functions in the predicted manner.As a result of such testing, it may be necessary to modify and refinethe prototype to obtain the desired results.

Once a satisfactory prototype is obtained, a rubber mold of theprototype is formed. Typically, this step involves encasing theprototype in molten rubber and allowing the rubber to harden so as toform a rubber mold of the prototype. The rubber encasing the prototypeis then split and the prototype is removed to yield an open rubber mold.

Once the rubber mold is obtained, it is used to form a wax casting. Thisstep typically involves pouring molten wax into the rubber mold andallowing the wax to harden to form a wax casting of the supportstructure.

After the wax casting is obtained, it is used to form a plaster mold.This step typically involves encasing the wax casting and plaster,allowing the plaster to harden around the wax casting so as to form aplaster mold.

The plaster mold can then be used to form a support structure.Specifically, molten bearing material, such as bronze, is poured intothe plaster mold so as to melt and displace the wax casting from themold. Thus, the plaster mold is filled with molten material and themelted wax is removed from the plaster mold. After the molten materialhardens, the plaster mold is removed from around the support structure.

As noted above, the first step in the investment casting method, indeedin any method, of producing support structures in accordance with thepresent invention is the production of a prototype bearing. Therelatively complex support structures of the present invention can beformed using simple manufacturing techniques.

With the foregoing in mind, it is believed sufficient to describe themethod of making a single support structure through the use ofelectrical discharge manufacturing and machining. A description of suchmanufacture demonstrates the ease with which the relatively complexsupport structure shapes of the present invention can be achieved.

Each support is initially in the form of a cylindrical blank having acylindrical bore as shown in FIGS. 5A and 5B. The blank is then machinedto provide a radial lubricating fluid groove as shown in FIGS. 6A and6B. For certain applications, it is desirable to further machine theblank to include facing grooves which are preferably symmetricallydisposed on the radial faces of the support structure as shown in FIGS.7A and 7B. The provision of such facing grooves ultimately results in asupport which is easily torsionally deflected. While the groove shown inFIGS. 7A and 7B are cylindrical, it is possible to provide taperedgrooves as shown in FIGS. 8A and 8B. This yields a support structurewhich exhibits improved deflection characteristics by virtue of theangled alignment of the support beams if it is preferable that thesupport beams converge at a point proximate the center line of theshaft. This ensures that flexibility occurs about the shaft center lineby establishing a center of action for the entire system such that thesupport may adjust to shaft misalignment. In essence, the tapering ofthe support beams causes the support to act in a manner similar to aspherical bearing by concentrating the support forces on a single pointabout which the shaft may pivot in all directions to correct anymisalignment. The arrows in FIG. 8A illustrate the lines of action ofthe deflection.

After the cylindrical blank is properly machined as shown in FIGS. 6Aand 6B, FIGS. 7A and 7B, or FIGS. 8A and 8B radial and/orcircumferential slits or grooves are formed along the radial face of themachined blank to define the support ring or support pads, the beamsupports and the housing. FIGS. 8C and 8D illustrate such grooves formedin the machined blank of FIGS. 8A and 8B.

When manufacturing low volumes of the support structures or prototypesof the support structures for use in the construction of a mold, thecuts or slits are preferably formed through electrical dischargemanufacturing or through the use of a laser. The machining of thecylindrical blanks to achieve the configurations illustrated in FIGS. 6Aand 6B, FIGS. 7A and 7B, FIGS. 8A and 8B or a similar shape can be donethrough conventional machine tools such as a lathe or the like.

The performance characteristics of the support structures of the presentinvention result from the relative shape, size, location and materialcharacteristics of the support member defined by the bores and cuts orslits formed in the machined blank. These parameters are largely definedby the dimensions and location of the radial circumferential bores, cutsor slits formed in the support structure in conjunction with the shapeof the machined blank in which the bores or slits are formed to yieldthe support structure.

While the construction of the support structures of the presentinvention is most easily understood by reference to the machiningprocess, larger quantities are preferably manufactured through theinvestment casting method of the present invention, and even largerscale production of the support structures contemplated by the presentinvention could be more economically performed through injectionmolding, casting, powdered metal, die casting, extrusion or the like.

In extruding a large number of support structures from a pipe-likecylindrical blank, radial lubricating fluid grooves, as shown in FIGS.6A and 6B can be provided along the length of the pipe-like cylindricalblank prior to extrusion. However, if facing grooves were desired in thesupport structure these can be individually defined after slicing theindividual support structure from the extruded and machined blank. Forthis reason, extrusion might not be a preferred method of producingsupport structure which require facing grooves to enhance torsionalflexibility.

Optimization of the support structure configuration for individualapplications precludes high stresses and insures long life. The absenceof moving parts eliminates pivot wear and durability concerns byeliminating the pivot pad contact stresses. Manufacturing tolerances arecompletely eliminated in all but the final support ring bore, thussimplifying the manufacturing process. EDM manufacturing is efficientand accurate for low volume applications, while high volume applicationscan be cast, molded, extruded or forged as discussed herein.

I claim:
 1. A fluid dampened support structurefor supporting a bearingfor at least radial movement, the support structure comprising a housingformed with a plurality of narrow cuts and grooves so as to define aplurality of circumferentially spaced beam members and a damping fluidfilling these spaces formed by the cuts and grooves such that thesupport structure is fluid dampened; wherein the support structurecomprises a plurality of circumferentially spaced pads each having twocircumferential ends, each pad being supported on the outer periphery ofthe support structure by a beam network which acts as a structuralspring, the beam network comprising a circumferential beam portionextending from each circumferential end of the pad and a radiallyextending beam portion extending between the circumferential beam andthe outer periphery of the housing, the beam portions together acting asa structural spring having a predetermined spring constant; means forchanging the spring constant of the structural spring after the springhas deflected a predetermined amount; means for adjusting thepredetermined amount of deflection which occurs before the spring ratechanges.
 2. The support structure of claim 1, wherein the supportstructure is formed of a single piece of material.
 3. The supportstructure of claim 1, wherein the support structure comprises aone-piece member formed with cuts and grooves to define a plurality ofbeams.
 4. The support structure of claim 1, wherein the supportstructure includes a thin annular membrane.
 5. The support structure ofclaim 4, wherein the membrane is dampened by an incompressible fluid. 6.The support structure of claim 1, wherein at least a portion of thesupport structure is integrally formed from a single piece of materialwith at least a portion of the bearing that it supports.
 7. The supportstructure of claim 1, wherein the support structure includes acylindrical bearing portion support surface for supporting the bearingportion.
 8. A damper for supporting a bearing, the damper comprising atleast three spaced circumferential pads for supporting the bearing and asupport structure including a structural spring supporting each of thepads for at least radial movement; wherein the structural spring has aspring rate that changes after a predetermined amount of deflection andfurther comprising means for adjusting the predetermined amount ofdeflection which occurs before the spring rate changes.
 9. The supportstructure of claim 1, wherein the fluid dampened support structureincludes a damper having an orifice configured such that the size of theorifice varies with deflection of the support structure.
 10. The supportstructure of claim 1, wherein the support structure comprises aplurality of circumferentially spaced pads each having twocircumferential ends, each pad being supported by a beam network on theouter periphery of the support structure, the beam network comprising acircumferential beam extending from each circumferential end of the padand a radially extending beam extending between the circumferential beamand the outer periphery of the housing.
 11. The support structure ofclaim 10, wherein the pads each include a damper portion extendingradially outward toward the outer periphery so as to define a thin gapbetween the damper portion and the outer periphery of the supportstructure.
 12. The support structure of claim 10, wherein each of thepads has an arcuate face having an inner diameter and wherein the innerdiameter of the arcuate face of each of the pads is such that thebearing being supported is in contact with the pads proximate thecircumferential center of the pads so that loads are applied in theproximity of the circumferential center of the pads.
 13. The supportstructure of claim 10, further comprising a plurality of deflectioncontrol members inserted into the unitary support structure to controldeflection of the beam network.
 14. A damper with adjustable dampingcharacteristics, for supporting a bearingthe damper for supporting thebearing comprising: a plurality of support pads for supporting thebearing, each of the support pads including a radially inner surfacesupporting the bearing, a radially outer surface and spacedcircumferential ends; a structural spring supporting each of thecircumferential ends of each support pad so as to allow at least radialmovement of the pads, each of the structural springs having apredetermined spring constant; and means for changing the springconstant of the structural spring after the spring has deflected apredetermined amount.
 15. The damper of claim 14, further comprisingmeans for adjusting the predetermined amount of deflection which occursbefore the spring rate changes.
 16. The damper of claim 14, furthercomprising a continuous generally circular base supporting each of thestructural springs, the base having a radially inner surface and aradially outer surface; the radially outer surface of each of the padsbeing spaced from the radially inner surface of the base so as to definea circumferentially extending gap, the gap being filled with anincompressible fluid so as to dampen movement of the pad with respect tothe base; and further comprising an orifice at each of thecircumferential ends of the gap to allow the incompressible fluid toflow into and out of the gap in response to movement of the pad relativeto the base, the orifice being defined by a radially inward extendingprotrusion formed on the radially inner surface of the base, the size oforifice being determined by the space between the circumferential edgeof the pad and the surface of the protrusion.
 17. The damper of claim16, wherein the size of the orifice varies with deflection of the padtoward or away from the base.
 18. The damper of claim 17, wherein theradially inward extending projection is tapered away from the pad suchthat the orifice becomes smaller with pad displacement.
 19. The damperof claim 17, wherein the projection tapers toward the pad so that theorifice increases with pad displacement.
 20. A damper structure fordamping forces applied to a bearing, the damping structure comprising aseries of spaced circumferential pads, each of the pads being supportedon a base member and including a radially inner surface, a radiallyouter surface and two circumferentially spaced ends; the radially innersurfaces of the pads supporting the bearing; a plurality of structuralsprings supporting the circumferential ends of the pads on the basemember, the structural springs being capable of deflection at leastradially so as to support the pads on the base member for movementtoward and away from the base; the radially inner surface of the padsbeing spaced from the base so as to define a damping gap having orificesproximate the circumferential ends of the pads, the damping gap beingfilled with an incompressible fluid such that movement of the padsrelative to the base is dampened by the incompressible fluid and causesthe incompressible fluid to flow into and out of the damping gap viaorifices at the circumferential ends of the pads.
 21. The damperstructure of claim 20, wherein the pads, springs and base are integrallyformed from a single piece of material.
 22. The damper structure ofclaim 20, further comprising means for varying the predetermined springconstant of the structural springs after a predetermined amount ofdeflection.
 23. The damper structure of claim 22, further comprisingmeans for varying the predetermined amount of deflection necessarybefore the spring rate is changed.
 24. The damper structure of claim 20,wherein the size of the orifice changes as the pad moves relative to thebase.
 25. The damper structure of claim 24, wherein the size of theorifice increases as the pad moves toward the base.
 26. The damperstructure of claim 25, wherein the orifice extends at an acute anglefrom the damping gap.
 27. The damper structure of claim 24, wherein thesize of the orifice decreases as the pad moves toward the base.
 28. Thedamper structure of claim 27, wherein the orifice extends at an obtuseangle from the damping gap.
 29. A damper for supporting a bearing, thedamper comprising at least three spaced circumferential pads forsupporting the bearing and a support structure including a structuralspring supporting each of the pads for at least radial movement furthercomprising a supply of fluid within the support structure such thatradial movement of the pads is dampened by the fluid.